RU2267042C2 - Power transmission device (variants) and continuously variable transmission - Google Patents

Abstract

FIELD: mechanical engineering, particularly continuous transmissions used as wide-range transmission mechanism.

SUBSTANCE: device for power transmission with variable velocity ratio comprises input 21 and output 27 members, leading draft member 7 and disc member 1 rotationally installed in case. Input member 21 transmits rotational moment of at least one leading draft member 7. Disc member 1 is coaxial to rotary shaft 2 brought into drive connection with at least one leading draft member 7 and is formed as driven member. Disc member 1 may be selectively moved in axial direction along rotary shaft 2 to change position of disc member 1 relative at least one leading draft member 7 to selectively change draft velocity ratio between leading draft member 7 and disc member 1. Output member 27 is displaced by rotary shaft 2 and is selectively brought into engagement with it to transmit rotational moment under determined velocity ratio relative input member 21.

EFFECT: extended velocity range.

27 cl, 4 dwg

Description

Technical field

The invention relates to power transmission systems, in particular to continuously variable transmissions containing a low inertia disk or wheel, which is driven by clutch rotors by means of a clutch. The invention can also be used as a wide-range transmission mechanism to expand the range of output speeds of the transmission system, which is engaged with a differential parallel mechanism with a disk and synchronized with it to control acceleration and deceleration at the output of the gearbox.

The level of technology.

Systems with a continuous (stepless) gear ratio change have been developed with the aim of creating a commercially competitive transmission that provides a wide selection of speeds in the continuous range. It is generally recognized that a continuous change in the gear ratio of the gearbox can provide a significant increase in the efficiency of the car engine, resulting in fuel economy. Continuously variable transmissions are also used in other technical applications to ensure continuous speed changes in a given range.

In cars with continuously variable gearboxes, power is transferred from the engine to the wheels through a gearbox with a variable gear ratio, which in turn must withstand high torque and satisfy other operational requirements. Several types of continuously variable transmissions have been developed, including traction transmissions and V-belt continuously variable transmissions based on belts and movable pulleys. Known systems have encountered significant problems, including rapid fatigue and failure of the main components, insufficient energy consumption, limited speed range, unstable gear ratio control, high cost and other problems that do not have an adequate solution that ensures widespread commercial use of such systems. Therefore, there is an urgent need for a stepless gearbox in which the listed problems would be solved and the production of which would be cost-effective.

SUMMARY OF THE INVENTION

This invention is a continuously variable transmission designed to change the gear ratio of the output shaft rotation speed relative to the input shaft driven by the vehicle engine or other energy source. In an embodiment of the present invention, the power transmission system includes a driven input piece engaged with at least one rotor for transmitting torque to the rotor. The disk part is mounted in articulation with the rotating shaft and in engagement with the rotor, so that the disk part is driven to transmit torque to the rotating shaft. The output part is driven by a rotating shaft, and torque is transmitted from the shaft to the output part, and the disk part can selectively move axially relative to the rotor and the rotating shaft, depending on the torque of the output load on the output part, thereby changing the torque, transmitted from the rotor. A gear mechanism driven by at least one rotor, which may include a planetary gear pair and a clutch, which have a parallel effect on the output torque of the disk part and the interacting rotating shaft, can be installed to expand the speed range of the output part and provide other advantages.

Thus, it is an object of the present invention to provide a power transmission system, in particular a continuously variable transmission, providing adequate energy consumption, a wide range of output speeds and efficient transmission ratio control in a robust and cost-efficient design.

A device for transmitting power with a change in the gear ratio of the speeds of the output part relative to the drive input part, according to the invention, comprises input and output parts mounted in a housing rotatably, the input part being in drive connection with at least one driving traction part for transmitting torque moment to said at least one leading traction part, a disk part located in articulation with a rotatably mounted shaft coaxially and in and connected to at least one driving traction part, this disk part is a driven at least one driving traction part for transmitting torque to the rotating shaft, the disk being selectively axially movable on this rotatably mounted shaft at least partially in response to the torque applied to the output part to change the position of the disk part relative to at least one driving traction part to selectively change the traction gear ratio between the leading traction part and the disk; moreover, the output part selectively engages with a rotatably mounted shaft to thereby transmit torque at a certain gear ratio of speeds relative to the input part.

Moreover, many leading traction parts can be located on the periphery of the disk part, each of the leading traction parts contains a conical rotor, the edge of which is located almost parallel and is in traction mesh with the disk part.

The power transmitting device may further comprise a shear mechanism associated with the disk part, for selectively moving the disk part to a predetermined axial position on a rotatably mounted shaft relative to at least one driving traction part to change the output torque transmitted to said rotatably mounted shaft.

In this device for transmitting power, at least one of said driving traction part may comprise a clamping mechanism designed to act on at least one leading traction part to selectively ensure effective traction contact of this leading traction part with a disk part.

In another embodiment, the device for transmitting power includes input and output parts mounted in a housing rotatably, the input part being in drive connection with at least one leading traction part for transmitting torque to this at least one leading traction part, a disk part mounted in conjunction with the rotating shaft coaxially and in a drive connection with the aforementioned at least one leading traction part, the disk part being driven at the edge with at least one leading traction part for transmitting torque to the rotating shaft, the disk can selectively move axially on the rotating shaft to change the position of the disk part relative to at least one leading traction part for selectively changing the traction gear ratio between the leading traction part and disk; while the output part is installed with the possibility of selective engagement with a rotating shaft to thereby transmit torque at a predetermined gear ratio of speeds relative to the input part, and at least one leading traction part is a rotor having a mounted shaft mounted in a mechanism bearings, and the clamping mechanism is a piston block mounted together with the bearing, the piston is mounted with the possibility of selective application of axial load to the rotor h through the shaft.

In the first embodiment, in the device for power transmission, at least one leading traction part may be a rotor containing a shaft, which at one input end contains a gear mating and driven with the input part, and at the other output end contains an output gear, moreover, the output gear is in a drive connection with the crown gear of the first set of planetary gears, and the set of planetary gears also includes a sun gear and a set of gears installed in the carrier, while the set of pl the gears are selectively engaged with the output part for transmitting torque.

In this case, the rotating shaft can be a driven disk part and is in a drive connection with an output sun gear connected in a drive connection with an output part, and the set of planetary gears is in a selective drive connection with a sun gear, so that a parallel differential drive system is provided through which the disk part and at least one leading traction part transmit torque to the output part with a certain gear ratio.

In a further embodiment, the power transmission device comprises input and output parts mounted in a housing rotatably, the input part being in drive connection with at least one leading traction part for transmitting torque to at least one leading traction parts, a disk part mounted in conjunction with a rotating shaft coaxially and in a drive connection with at least one leading traction part, wherein the disk part is driven by at least one driving the traction piece for the transmission of torque to the rotating shaft, the disc may be selectively moved axially on a rotating shaft for changing the position of the disc parts relative to at least one leading traction parts for sampling the torque transmission ratio between the driving of the traction piece and the disk; moreover, the output part selectively engages with the rotating shaft to thereby transmit torque at a predetermined gear ratio of speeds relative to the input part, and the disk part contains a sleeve mounted on a rotating shaft, the sleeve is torsionally connected to the rotating shaft by a screw groove made in the rotating shaft .

In a first embodiment, the power transmission device may further comprise a differential gear unit operable in connection with the input power source, wherein the differential gear unit comprises a ring gear selectively engaged to drive a set of output planet gears for selectively driving output shaft.

Moreover, in the device for transmitting power, the ratio of the output torque transmitted to the output part from the rotating shaft or from the differential gear unit, which are both driven by at least one leading traction part, can be matched at the synchronization point of the ratio based on the required output torque from the output part due to the axial movement of the disk part.

In addition, the device for transmitting power may further comprise a control system associated with the differential gear unit, the control system receives control signals related to at least the input and output transmission speeds, while the control system interacts with the differential gear unit to provide required output speed and torque, selected to match the output power of the gearbox with the input power of the power source.

The continuously variable transmission according to the invention comprises a disk part mounted in conjunction with the rotating shaft coaxially, the disk part being in drive connection with a plurality of driving traction parts, a plurality of driving traction parts including a leading surface parallel to the axis of the disk part, the disk part being selectively mounted axial movement on a rotating shaft to change the position of the disk part relative to the set of leading traction parts To selectively change the traction gear ratio between a plurality of driving traction parts and a disk part, a plurality of driving traction parts are set so as to be driven by an input power source from one end thereof, the second end being in a drive connection with a gear system comprising a ring gear in leading engagement with a set of output planetary gears, while the rotating shaft is mounted to drive the output sun gear associated with the set of gears planetary gears, while the sun gear is also in a drive connection with a set of planetary gears to provide a parallel differential drive through which the disk part and a plurality of driving traction parts transmit the output torque to the output part connected in the drive connection to the planet gears.

In this embodiment, the continuously variable transmission may further comprise a shear mechanism for controlling the axial position of the disk part relative to the plurality of driving traction parts, whereby the gear ratio between the disk part and the driving traction parts is changed in order to control the output speed and torque of the output shaft.

In another embodiment, the continuously variable transmission comprises an input power source coupled in a drive connection to at least one leading traction part, this at least one leading traction part having a leading surface, a disk part driven by at least one the leading traction part, and this disk part is torsion connected with the coaxial shaft and mounted with the possibility of selective movement relative to the leading surface to change the gear ratio of the torque transmitted about it from at least one leading traction part, a differential gear unit connected to an input power source, the differential gear unit comprising a ring gear selectively engaged to drive a set of output planet gears for selective drive in the movement of the output sun gear, shaft, while the coaxial shaft is designed to transmit torque to the output sun gear with a predetermined gear ratio together with curled differential gear unit, and the position of the said disk parts relative to the leading surface is selectively adjusted to change the gear ratio of torque and speed transmitted to the sun gear by said coaxial shaft to the differential gear unit.

In this case, the continuously variable transmission may further comprise a shift mechanism for controlling the axial position of the disk part relative to the leading surface.

In addition, in a continuously variable transmission, the position of the disk part relative to the driving surface can be changed in response to the output torque to selectively maintain a synchronized relationship between speed and torque transmitted to the sun gear by the coaxial shaft relative to the differential gear unit.

The continuously variable transmission may also further comprise a control system associated with the shift mechanism and a differential gear unit, the control system controls the shift mechanism to selectively change the position of the disk part, this control system also receives control signals regarding at least the input and output transmission speeds, the control system controls the matching of the output power of the transmission with the input power source.

In addition, in a continuously variable transmission, the differential gear unit may comprise a gear system comprising at least two sets of planetary gears providing a predetermined gear ratio to provide the desired speed range at the output.

In a continuously variable transmission, at least one leading traction part also comprises a clamping mechanism designed to act on at least one leading traction part to selectively clamp this at least one leading traction part in order to ensure effective traction contact with disk part.

In continuously variable transmission, the gear system may be a selectable multi-range gear system.

In a continuously variable transmission, the shear mechanism may also comprise a helical groove made in a rotatably mounted shaft.

In the device for transmitting power, the shear mechanism may include a helical groove made in a rotatably mounted shaft.

In a continuously variable transmission, the shear mechanism may comprise a helical groove formed in a coaxial shaft.

The power transmission device may further comprise a control system associated with the differential gear unit and the disk part, the control system responding to control signals relative to the axial position of the disk part, while the control system interacts with the differential gear unit to provide the required selected gear ratios of the speeds between the disk part and differential gear unit.

In a device for transmitting power, the control system may respond, at least in part, to control signals related to the axial position of the disk part.

The continuously variable transmission may further comprise a control system associated with the gear system and the disk part, while the control system responds to control signals that at least partially relate to the axial position of the disk part, the control system interacting with the gear system to provide the required selected gear speed relationships between the disk part and the gear system.

The continuously variable transmission may further comprise a control system associated with the differential gear unit and the disk part, the control system responding to control signals relative to the axial position of the disk part, while the control system interacts with the differential gear block to provide the required selected gear ratios of the speeds between the disk part and differential gear unit.

A brief description of the drawings.

Figure 1 is a side view in partial section in the longitudinal direction of a mechanism made in accordance with this invention.

FIG. 2 is a diagram for controlling fluid pressure for the mechanism shown in FIG. 1.

Figure 3 is a diagram of the operation of the clutch and valve system controlling the shift for the mechanism shown in Figure 1.

Figure 4 is a separate section of a conical block from the mechanism shown in Figure 1.

Information confirming the possibility of carrying out the invention.

Figure 1 shows a preferred embodiment of the invention in the transmission of a vehicle (vehicle) for transmitting torque from the engine to the drive wheels of the vehicle. The transmission may include a central driven traction part or a disk 1 mounted coaxially on the shaft 2. It should be noted that all the details of the mechanism are inside the casing 3. We also note that the shaft 2 is mounted with appropriate bearings, such as a set of bearings 4 on the front end of the shaft 2 , which can be thrust bearings to prevent axial movement of the shaft 2. The disk 1 is rigidly mounted on a concentric sleeve 5, which can be made in the form of a recirculating ball bearing key, corresponding in shape me and included in the spiral ball groove of the shaft 2. The sleeve 5 is arranged in such a way that the disk 1 can move along the axis of the shaft 2 by means of mutual friction-type screw movement while simultaneously transmitting torque. The structure of the ball key of the sleeve 5 and the groove of the shaft 2 is well known and therefore not shown in detail.

The positioning lever 6 is connected to the sleeve 5 by a corresponding bearing, such as a thrust bearing in the lever 6. The lever 6 is designed to move along the axis of the shaft 2 and thereby control the axial position of the disk 1, while allowing free rotation of the disk 1 and shaft 2.

A plurality of conical rotors 7 are symmetrically mounted on the periphery of the disk 1 so that the inner faces of the cones 7 are parallel to the shaft 2 and are in frictional engagement with the edge of the disk 1. In a preferred embodiment, eight cones 7 are mounted on the periphery of the disk 1, although only two cones are shown 7. For traction parts, such as disk 1, the service life of the traction element due to material fatigue, as well as the bearings in which they can be installed, is calculated according to the following known form Ole:

It follows that the applied load should be relatively small in order to prevent the rapid failure of the part due to material fatigue. In addition, the applied load has a much greater impact on the life of the part than the speed of rotation of the elements. The output torque of the disk 1 is also proportional to the applied contact traction load and the number of traction points of contact. On the other hand, the optimal number of driving traction elements or cones 7 can be selected based on the operating conditions and purpose of the power transmission system. In a preferred embodiment of the present invention, the maximum number of contact points is used in order to provide maximum output power at the output rotation speed determined by the differential mechanism unit described below. The disk part 1, as well as the traction drive part 7, is designed and constructed in such a way as to withstand extremely high rotation speeds without cases of rapid failure. Note also that the cones 7 in this embodiment are cones with an angle of fifteen degrees (15 °) at the apex and a ratio of diameters of about three to one (3: 1), although other embodiments may contain cones with different angles at the apex, sizes and diameters. 4, a cone block is shown in more detail. Each cone 7 contains a concentric shaft, which is mounted at the ends using bearings 8 and 9. The pinion gear 10 is mounted on the front end, and the output gear 11 is mounted on the rear end of the shaft, the front end of the shaft is mounted on the thrust bearing 12, as shown in the drawing. Bearings 8 and 9 are preferably needle roller bearings capable of withstanding extremely high rotational speeds and relatively large radial loads. Bearings 8 and 9 are mounted in guide blocks 13 and 14, respectively, which, in turn, are attached to the cover 15 of the cone block attached to the main casing 3. The concentric piston 16 abuts the thrust bearing 12 from the back and can be sealed with rings 17 inside the reinforced cylinder 18. The cylinder 18, in turn, abuts against the edge of the cover 15 so that the piston 16 creates axial thrust through the bearing 12 on the cone 7, while the fluid pressure is directed through the inlet 19 into the cylinder 18 and into the piston 16. Cylinder 18 and piston 16 setting extend in a fixed position near the rotating block of the cone 7 with a stop through the bearing 12. We also note that the bearings 8 and 9 and the axlebox of the shaft of the cone 7 are mounted so as to allow a small but adequate axial movement of the cone 7. The pinion gear 10 can be mounted on the shaft and located on this shaft due to the force directed through the bearing 12. The output gear 11 can also be mounted and held on the shaft by means of a locking ring 20, shown in Fig.4. Both gears 10 and 11 are mounted on the shaft in the usual way.

As shown in FIG. 1, the input ring gear 21 is normally engaged with the pinion gear 10 and is mounted with a set of bearings 22 in conjunction with a front mounted cover 23. Bearings 22 are preferably used which are resistant to radial and axial loads and are fixed in the cover 23 lock cover 24. A set of bearings 4 prevents axial movement of the shaft 2 and is anchored concentrically together with the gear 21 by the stopper 25. The gear 21 is designed for communication and is driven accordingly a traveling shaft, for example, an engine flywheel, or, as shown in this embodiment, coupled to a conventional input shaft 26 driven by any suitable power source, such as an electric motor, gas or steam turbine, automobile engine, or the like. It should be noted that in the described embodiment, the gear 21 can be driven by the engine of a truck or car.

The output ring gear 27 is in normal engagement with the output gears 11 and is intended to drive a planetary mechanism, which is preferable to use a multi-band gear and a clutch system in parallel operation at the output torque together with the disk 1 and the shaft 2. As shown in FIG. 1, the sleeve or neck of the gear 27 is mounted in the respective bearings.

The multi-range gear system may include a planetary gear carrier 28 mounted concentrically and rotatably on the neck of the gear 27 due to the bearing, as shown in FIG. A plurality of planetary gears 29 are installed on the carrier 28, of which there must be at least four, and which are mounted in normal engagement with the corresponding sun gear 30. The sun gear 30 is concentrically attached to the neck of the gear 27 and engages in conventional engagement with the planet gears 29. Crown a gear 31 with internal teeth, which may be an integral part of the clutch housing 32, is made in such a way that it engages normally with the outer side of the planet gears 29, opposite the sun gear 30. The outer edge of the carrier 28 has the shape of a clutch drum, and the size is suitable for friction engagement with the clutch belt 33. The shape and dimensions of the outer edge and the clutch drum of the carrier 28 must also satisfy the adhesion conditions between the ring gear 31 and the planet gears 29 and must allow their free rotation. To fix the clutch tape 33 on the carrier 28, a conventional mechanism is used.

Clutch housing 32 may include a conventional multi-plate clutch assembly for engaging and disengaging the clutch between the clutch housing 32 and the drive flange 34 mounted on the neck of the gear 27 along the side of the sun gear 30. The clutch housing 32 has a neck that extends the housing into axial direction from the back to mate with the bearing support and to improve articulation with other elements. The secondary sun gear 35 can be installed inside the clutch housing 32 and has an elongation in the rear direction for interfacing with an additional bearing support. The sun gear 35 has teeth that normally engage and drive a second set of planet gears 36 mounted in the carrier 37 of the secondary planet gear. The carrier 37 is supported by suitable bearings mounted on the clutch housing 32 and the sun gear 35. The secondary ring gear 38 with internal teeth, which is an integral part of the clutch housing 39, engages with planet gears 36 on the opposite side of the sun gear 35. The secondary housing the clutch 39 is mounted rotatably on the housing of the clutch 32 using an appropriate set of bearings.

In the same way as the clutch housing 32, the secondary clutch housing 39 includes a multi-plate clutch unit for turning the drive flange mounted or integral part of the clutch housing 32 on and off. The pressure clutch 40 is used conventionally, and it Designed to provide selective fluid pressure on clutches 32 and 39, respectively. The secondary clutch belt 41 is also present and installed to provide frictional adhesion to the corresponding surface of the outer edge of the clutch housing 39.

A set of double-length planet gears 42 can be rotatably mounted on carrier 37 such that one end engages planet gears 36 without contacting the sun gear 35 or ring gear 38. For example, three gears 36 and three gears 42 can be mounted symmetrically in the desired configuration. We also note that, being meshed with gears 36, gears of double length 42 have a direction of rotation opposite to the direction of rotation of gears 36. The rear end of gears of double length 42 is engaged with a ring gear 43 with internal teeth, which has an integral connection with the output crown gear 44, both gears being mounted in respective bearings.

The output ring gear 44 has internal teeth and normally engages with a set of output planet gears 45. The output planet gears, of which there must be at least four, are distributed symmetrically and mounted on coaxially mounted spindles made integral with the output sleeve 46, which is integral with the output shaft 47 installed in the corresponding set of bearings 48. The output planet gears 45, as well as other sets of planet gears, in the preferred embodiment, o existences are equipped with needle roller bearings to provide low friction rotation. The planet gears 45 are in normal engagement with the output sun gear 49, which is mounted on the rear end of the shaft 2. Note that the gear 27 supporting (bearing) 32 and gear 35 are arranged so as to allow axial installation of the shaft 2 if necessary.

To control the positioning lever 6 on the casing 3, a pair of coaxial control cylinders 50 is installed on the opposite side with respect to the multi-range gear system described above. Note that for simplicity, only one proximal cylinder 50 is shown in FIG. 1. Each of the cylinders 50 comprises a rod 51 of a corresponding length and stiffness, which is rigidly fixed to the corresponding external end of the lever 6 so as to effectively control the axial position of the lever 6 due to the operation of the cylinders 50 and rods 51. The cylinders 50 have conventional openings 52 and 53 for connection to a fluid pressure source (not shown). Another suitable mechanism may also be used to move the lever, not just the hydraulic control unit shown here.

Figures 2 and 3 schematically illustrate the operation of a system controlling a cylinder 50, clutches 32, 33, 39, 41 and all of the pistons 16. The control system in a preferred embodiment is a hydraulic pump 54 that can be driven by the same conventional energy source which drives the input shaft 26, and which provides the pressure of the liquid, preferably oil-based lubricant, in the main line 55. The pump 54 draws liquid through the inlet pipe 56 from the oil sump 57, which, although not shown n in the figure, is mounted to the lower part of the casing 3 in the usual way. Pump 54 may be a variable displacement constant pressure pump as shown in the figure, or may be any other type suitable for a given application. Other types of non-hydraulic control systems are also known, and they have also been considered in the context of the present invention.

The line 55 in the hydraulic system transfers the main pressure to the control module 58, feedback module 59, pressure relief valve 60, pressure relief valve 61, pressure relief valve 62, start valve 63, pulse valve 64, as well as other elements that may be required supply of such pressure. Note that the pressure relief valves 60, 61, and 62 are controlled by the pilot pressure in the common line 65 created by the feedback module 59.

The pressure relief valve 60 directs the controlled pressure to line 66 and through the control valve 67 and line 68 to the directional valve 69 and to the pressure-controlled pressure relief valve 70. The pressure relief valve 70 is controlled by the pilot pressure from line 66. The control valve 67 is installed to allow free flow from the line 66 to line 68, but prevents backflow. Another control valve 71 is connected to line 66 and is designed to provide free flow from the oil pan 57 to line 66 in case of negative pressure, but stops the return flow to the oil pan 57. The pressure sensor 74 is also connected to line 68 to control the pressure in this lines.

The directional valve 69 may be a pilot-operated, four-way valve with four directions and is connected via lines 72 and 73 to the openings 52 and 53 of the cylinder 50, respectively. Line 75 is the discharge line from valve 69 to oil pan 57. Note that when valve 69 is in the middle position, fluid may flow between lines 72 and 73 and pressure from line 68 is stopped. To switch the valve 69 to the forward or reverse position, an alternating pilot pressure is applied from the control module 58 through lines 76 and 77. A pilot pressure control valve 78 is installed on lines 76 and 77 so that, if necessary, the pilot pressure can be disconnected from valve 69. For example , in the car the PARKING - NEUTRAL - MOTION switch is usually set, then it is necessary to disconnect the pilot pressure from the valve 69 in the PARKING or NEUTRAL positions, and turn on the pilot pressure when the switch is in position MOVEMENT. A pilot pressure line 79 may be used to connect the valve 78 to the PARKING - NEUTRAL - MOTION switch, which is not shown here, in order to control valve 78.

The pilot pressure is routed through one of the lines 76 or 77, selectively, by the control module 58 so that the pressure in line 77 turns on the valve 69 in the forward direction and the pressure in line 76 turns on the valve 69 in the opposite direction. Note that the pilot pressure line 76 is also connected to the pilot pressure reversing valve 80 shown in FIG. 3, the operation of which is discussed below. When the valve 69 is turned on in the forward direction, pressure is supplied from line 68 to the line 73 in order to provide the necessary buoyancy force on the cylinder 50, and line 72 connects the valve 69 to the discharge line 75. On the other hand, when the valve 69 is turned on in the opposite direction , pressure is supplied from line 68 to line 72 to provide a pulling force on cylinder 50 to select the position of disk 1, and line 73 is connected via valve 69 to exhaust line 75.

In order to reduce the resistance to fluid flow during the rapid movement of the cylinder 50 between lines 72 and 73, a pair of pressure-controlled bypass valves and a combination of control valves 81 and 82 in close proximity to the cylinder 50 can be turned on so that they are controlled by pilot pressure through lines 83 and 84, respectively. A detailed description of the operation of valves 81 and 82 is given below.

The pressure relief valve 61 directs the controlled pressure through line 85 to all channels 19 and, thereby, to all the pistons 16, as shown in FIG. 1 and FIG. 4. Note that the channels 19 can be placed in the axial direction, as shown in FIG. 1, or in the radial direction, as shown in FIG. 4, depending on which is more suitable for a particular application. For the same connection of line 85 and channels 19, a board with a common collector can be mounted. As shown in FIG. 3, line 85 is also connected to other elements of the control system.

The pressure relief valve 62 directs pressure through line 86 to the shift valve of the clutch 87, as shown in FIGS. 2 and 3, for controlled communication with the respective clutches of the mechanism (FIG. 1).

As shown in FIG. 2, control module 58 may be controlled by actuator 88 and communicate with feedback module 59 through pilot pressure via line 89. Alternatively, module 58 and module 59 may be combined into a single module. As the actuator 88, any suitable mechanism may be used, such as a pedal, a lever, a remote control mechanism, or any other means for controlling the module 58.

It is preferable to provide the control system with control signals based on the functioning parameters of the mechanism, which may include, for example, a tachometer for measuring the input speed 90, as shown in FIG. 2, connected to the input power source and providing a feedback signal at the input speed to module 59 along line 91 In the same way, the output speed control signal from the tachometer 92, which can be connected to the output shaft 47, as shown in FIG. 1, can be supplied to module 59 via line 93. Line 93 can also be connected to speed control module and throttle valve 94. Another line 95 connects module 58 to module 94. Line 73 is also connected to module 94.

In a preferred embodiment, module 94 functions as a engine speed control module in accordance with a signal from module 58 via line 95 and provides a throttle position signal via line 93 to module 59. To limit engine speed when operating at a low gear ratio, via line 73 to module 94 is supplied with a pressure signal. Module 94, if necessary, can of course be combined with module 58 and / or module 59.

Consider the operation of the control system in conjunction with the gearbox (Figure 3). The activation of the start valve 63 is shown in the drawing for a general illustration and introduction of the plunger 96 and the engine 97 connected thereto. The engine 97 has contact surfaces 98 and 99 formed so as to provide effective contact with the lever 6. As an alternative, the valve 63 can mounted remotely from the lever 6 and turned on by an electric actuator, rather than the engine 97 moved by the lever 6. Line 55 brings the operating pressure to the valve 63, and line 100 is a drain. Via lines 101 and 102, a downward or upward shear pressure is supplied from the valve 63, respectively, to the control valve 103 and to the shift valve of the clutch 87.

The control valve 103 receives a working pressure proportional to the traction pressure on the pistons 16 through line 85, and also receives a temporary lag signal through line 91, which comes from the engine tachometer 90. Valve 103 has a drain line 104 and three different working output lines: 105, which directs the pressure reduction signal to the pressure relief valve 62; a line 106 that directs the pressure reduction signal through the valve 80 to the corresponding line 107 and thereby to the pressure relief valve 60; and a line 108 that directs the pressure reduction signal through the valve 80 to the corresponding line 109 for the pressure relief valve 60.

The shift valve of the clutch 87 receives the control pressure through the line 86 for actuating the clutch, and also has a drain line 110. Lines 111, 112, 113 and 114 direct the pressure of the actuation from the valve 87 to the clutches 32, 33, 39 and 41 respectively. The method of connecting the lines 111, 112, 113 and 114 with the respective clutches and the clutch actuator are conventional and therefore not shown here.

The pulse valve 64, in addition to the other lines discussed above and shown in FIG. 3, has a drain line 116. The lines 107 and 109 are connected through the valve 64 to provide the desired action inside the valve 64. Line 115 sends a switching pulse to the valve 87 as soon as such an impulse is generated by the valve 64. The impulse valve 64 is configured to provide an impulse to line 115 to valve 87 as soon as pressure appears in line 107 or as soon as the applied pressure is released through line 109.

Note that all valves and the corresponding control circuits intended for them can be manufactured together in one valve block, which in turn can be compactly mounted in the oil pan of the device shown in FIG. 1. Moreover, although the control system in this embodiment is substantially hydromechanical, any or all of the elements of the control system may be electrical or electronic. In addition, if the maximum improvement of the control system is required, then a microcomputer or microprocessor can be included in its composition to take into account factors such as input speed and torque, output speed and torque, oil temperature and viscosity, clutch slip moment, and others important operating parameters or environmental parameters, and thus continuously control and optimize the operation of the device as a whole.

And, on the contrary, in applications that do not require the maximum improvement of the control system, a much simpler control system is suitable. For example, control valve 103, pilot pressure reversing valve 80, and pulse valve 64, as will be shown below, are necessary only to provide soft or imperceptible switching jumps, so they can be eliminated. Also in applications where the input speed is constant or independently controlled, the speed control module 94 and the feedback module 59 can be omitted.

The operation of the device will be considered for this embodiment, which can be used in trucks or cars, when the ring gear 21 is driven by the engine of a truck or car. In operation, gears 10 and therefore the cones 7 are driven by the ring gear 21 and have relatively high rotational speeds depending on the gear ratio between the gear 21 and gears 10. For example, at a maximum speed of the engine shaft of 4000 revolutions per minute, the cones 7 are preferably an embodiment will have a rotation speed of up to 20,000 rpm. Obviously, variations are possible over a wide range of speed and gear ratio. Also, the gears 11 drive the output gear 27 at a predetermined speed, which, in turn, drives the output ring gear 44 through a multi-range gear system at a selected gear ratio.

At this time, the pump 54 provides the necessary pressure, as described above, so that the pressure relief valve 61 directs at least the set minimum pressure through line 85 to the pistons 16. Thus, sufficient force is applied to the cones 7, so that the cones 7 enter into the same and adequate traction contact with the edge of the disk 1. Therefore, the disk 1 and the shaft 2 rotate in the same direction as the gear 21, and with a relative speed determined by the axial position of the disk 1 relative to the cones 7. The shaft 2 drives the solar shhe stubble 49.

The operation of the gearbox will be considered in relation to an embodiment in an operator-driven vehicle. The actuator 88 may initially be in a neutral position, so that the module 58 does not transmit any signal through line 89, but at the same time transmits a minimum signal through line 95 towards the module 94 to maintain only the minimum operating speed of the engine. In the absence of a signal through line 98, module 59 is inactive and does not transmit any pressure control signal through line 65, regardless of the speed signals coming through lines 91 or 93 from tachometers 90 and 92. In neutral mode, module 58 transmits pilot forward direction pressure through the line 77, however, valve 78 is initially turned off and blocks lines 76 and 77 when the normal PARKING - NEUTRAL - MOTION switch is in the PARKING or NEUTRAL positions. In these positions, the valve 69 is initially in the neutral position and the cylinder 50 can move freely, that is, no pressure is applied to it.

When operating in a small speed range, the clutches 33 and 39 are selected by valve 87 and the actuating pressure is applied to them from the pressure relief valve 62. In neutral mode, in the absence of pilot pressure through line 65, the pressure relief valve 62 is preset to the position in which it sends to all clutches only the minimum pressure to actuate them, at which they are only in easy engagement. In the low-speed range, the clutch 33 holds the carrier 28 in a stationary position, so that the planet gears 29 rotate on stationary axes and in the opposite direction with respect to the direction of rotation of the driving sun gear 30. The crown gear 31 also rotates in the opposite direction with the planet gears 29. B In a preferred embodiment, the ring gear 31 may have a diameter twice that of the sun gear 30, therefore, in the low speed mode, the ring gear 31 and therefore the clutch housing clutch 32 rotate in the reverse direction at half speed. In addition, since the clutch 39 is engaged when operating in the low speed range, the clutch housing 32, the clutch housing 39, the ring gear 38, the planet gears 36, the carrier 37, the planet gears 42, the ring gear 43 and the output ring gear 44 all rotate together as a unit, and thus the output ring gear 44 rotates in the opposite direction at half speed while operating at a low speed range.

It should be noted that the sun gear 49 always rotates in the forward direction due to the pulling force between the cones 7 and the disk 1. Therefore, when the disk 1 is at the “synchronous gear ratio point” relative to the cones 7, only in the low speed range, the rotation speed in the forward the direction of the gear 49 corresponds to the speed of rotation of the gear 44 in the opposite direction, so that the planet gears 45 rotate in stationary positions and no output torque is transmitted either to the sleeve 46 or to the shaft 47. Such conditions p Works are commonly referred to as "gear neutral." It should also be taken into account that any positive value of the torque between the disk 1 and the shaft 2 affects the sleeve 5, which, together with the disk 1, receives axial displacement along the shaft 2 with the help of a screw key in the direction of the narrow ends of the cones 7, reducing the traction gear the ratio and speed of rotation of the disk 1 until the mentioned “synchronous point” is reached and the torque becomes zero. Similarly, any negative value of the torque between the disk 1 and the shaft 2 affects the sleeve 5, which together with the disk 1 receives axial displacement along the shaft 2 with a screw key in the direction of the wide ends of the cones 7, increasing the traction gear ratio and the speed of rotation of the disk 1 until a "synchronous point" is reached and the torque becomes zero. Thus, the disk 1 moves to the position of the synchronous point and remains in it due to the reaction to the torque through the shaft 2 and the sleeve 5, so that the system is self-synchronizing.

In the low speed range, when the PARKING - NEUTRAL - MOTION switch is turned to the MOTION position and pressing the actuator 88 in the forward direction relative to the neutral in order to provide the output torque on the shaft 47 in the forward direction, the signal is turned on through line 89 to activate module 59, it turns on valve 78, valve 69 is biased forward and a proportional signal is transmitted through line 95, so module 94 requires a proportional increase in engine speed. As the engine speed increases, the speed signal from the tachometer 90 enters line 91 and causes the module 59 to transmit a proportional pressure increase signal to line 65. Thus, the pressure relief valve 60 directs the increased pressure to cylinder 50, which causes the rod 51 to pop out. with proportional force. As a result, the lever 6 and, therefore, the disk 1 are moved towards the wide ends of the cones 7, increasing the traction gear ratio and increasing the relative speed of rotation of the sun gear 49. Thus, the rotating planet gears 45 begin to rotate in the forward direction, which leads to the rotation of the shaft 47 in the forward direction.

In this mode of operation, any reaction to a positive torque from shaft 47 backward through shaft 2 causes the sleeve 5 to move in the opposite direction with an increase in the force on the rod 51, so that the actuating pressure in the cylinder 50 will be proportional to the output torque of the shaft 47. Note also that sensor 74 can detect both torque and pressure.

Further operation of the actuator 88 causes the module 94 to require an increase in the rotational speed of the engine and engine tachometer 90 through line 91, as a result of which the module 59 transmits a pressure increase signal to line 65, so that the pressure relief valve 60 directs the increased pressure to the cylinder 50. Thus, the output torque on the shaft 47 is increased since its output torque is proportional to the bias pressure in the cylinder 50.

During this time, always proportional to the output torque and in response to the pressure signal on line 65, the pressure relief valve 61 directs the proportional pressure through line 85 to the pistons 16 to maintain the desired or sufficient traction between the cones 7 and the disc 1 to prevent slippage between them, at any permissible values of speed and torque at the load. Note that since the cones 7 are preferably mounted symmetrically around the disk 1 and all the cones 7 have the same frictional contact stress with the disk 1, no net radial load is applied to the disk 1. The effective traction between the cones 7 and the disk 1 is ensured by an elastohydrodynamic oil film on mating surfaces and between points of contact at all permissible values of speed and load. Such elastohydrodynamic adhesion is achieved through the use of various lubricants belonging to the class of “fluid adhesion”, well known in the prior art. At the same time, the coefficient of adhesion can reach up to nine percent (9%) and higher. The cones 7 and the disc 1 are preferably made of materials such as bearing steel, which can withstand constant high contact pressure at high operating speed. Methods of applying lubricants to the traction surfaces of the cones 7 and the disc 1, as well as to bearings, gears and clutches, are well known.

During operation, the torque of the disk 1 transmitted through the sun gear 49 to the shaft 47 is proportional to the gear ratio between the gear 49 and the sleeve 46. In this embodiment, the gear ratio between the gear 49 and the sleeve 46 is four to one (4: 1), and the gear ratio between the ring gear 44 and the sleeve 46 is four to three (4: 3). Therefore, in this embodiment, the disk 1 transfers one fourth (1/4) of the torque to the shaft 47, while the ring gear 44 and its drive gear and clutch system transfers three quarters (3/4) of the torque to the shaft 47. Of course, in other embodiments, other combinations of gear ratios are possible.

Also during operation, it is always proportional to the output torque and in response to the pressure signal on line 65, the pressure relief valve 62 directs the proportional pressure through line 86 to the respective clutches to maintain sufficient force on the clutches to prevent them from slipping at any permissible speed and load moment.

In a preferred embodiment, the control system operates in such a way that the output speed signal from the tachometer 92 is compared with the input speed signal from the tachometer of the engine 90, so that the rotation speed of the output shaft 47 increases in proportion to the increase in the input speed for a given output torque value and bias pressure. Thus, being pre-calibrated, the module 59 compares the corresponding signals from the tachometers 90 and 92 so that the input power is consistent with the output power in accordance with the equation power = torque × revolutions / minute. With this method, the engine can never be overloaded due to the torque on the load, instead, the engine speed increases to match it with the required power on the output shaft 47. Accordingly, with a low speed of rotation of the output shaft 47 and this torque on the load, it is enough relative low engine speed to match power output. With a higher output speed and a given torque on the load, a higher engine speed is required, while with a higher output speed, but a lower value of the load torque, it is necessary to reduce the engine speed. In the same way, with a low output speed but a higher load torque, a relatively higher engine speed is required. The possibility of continuously varying the gear ratio provided by the invention facilitates continuous power matching, as described above, by appropriately calibrating the module 59, any given value of the percentage of engine load can be maintained at any possible output speed and load moment.

Module 94 can be arranged in such a way that upon reaching the permissible maximum value of the output torque represented by the shear pressure in line 73, the operation of module 94 limits the engine speed to a value sufficient for power matching, regardless of the degree of pressing the actuator-pedal 88. In addition, when the engine reaches full load, which is reflected in the complete opening of the damper, module 94 transmits the corresponding signal through line 93, reducing the signal on line 91, so that modulus value of the signal 59 reduces the pressure in line 65, thereby reducing the output torque to the limit set by the calibration unit 94. Thus, the gear ratio of the transmission decreases, allowing the engine to increase the power delivered to the load. This sequence of downward switching actions is performed whenever the output load increases, when the pedal 88 is pressed further to increase the output speed.

When operating in the low speed range, when sufficient displacement pressure is maintained in the cylinder 50 and sufficient resulting torque is maintained, the disk 1 is shifted towards the wide ends of the cones 7 and the speed of the output shaft 47 is increased until the disk 1 reaches the ends of the cones 7, which is the limit of its forward movement. As shown in figure 3, the lever 6 will be in position 6-U, in which it comes into contact with the surface 98 of the engine 97, thereby shifting the start valve 63 to the upper position corresponding to a higher operating range. In the upper position mode, valve 63 directs pressure through line 102 to control valve 103 and clutch switch valve 87, setting both valves to switch up the clutch. Once the control valve 103 is set in any of the up or down switch modes, pressure is directed from line 85 through line 105 to pressure relief valve 62, which in turn reduces pressure in line 86 to the respective clutches, so that active clutches can slip with high torque in the load. This is done to prevent ripple of the torque on the shaft 47, until the disk 1 is not resynchronized, as will be described below. In addition, as soon as the control valve 103 is set upward, pressure is directed from line 85 through line 106, through valve 80, through valve 64 and line 107 to pressure relief valve 60, which in turn reduces the bias pressure on cylinder 50 for in order to compensate for the inertia during resynchronization of the disk 1, as will be described below.

When pressure arises in line 107, valve 64 provides a pulse through line 115 to valve 87. Since valve 87 is now set to switch upward, the pulse from valve 64 steps the valve 87 forward so as to select the next combination of a higher clutch range. Thus, when switching up from the lower to the next range, the clutch 41 engages, the clutch 39 remains engaged and the clutch 33 disengages. Then, clutch 39, clutch housing 32, carrier 37, planet gears 36, planet gears 42, sun gear 35, ring gear 43 and ring gear 44 are stopped by clutch 41 and remain in this mode until they move to the next operating range. Clutches 31 and 33 are in free rotation.

Whenever the ring gear 44 is switched from the reverse rotation mode to the stop mode when switching from the lower to the next range at a given shaft speed 47, the reaction of planetary gears rotation 45 leads to a decrease in the torque applied through the sun gear 49 to the shaft 2 and through the sleeve 5 to the disk 1, thereby causing a force to move the disk 1 towards the narrow ends of the cones 7 to a new synchronization position. This process is called resynchronization. During resynchronization, the active clutches are adjusted as described above to allow the clutches to slip so that the output torque on the shaft 47 does not change during resynchronization. After the resynchronization process is completed, the disk 1 is in the new synchronization position closer to the narrow ends of the cones 7, and the rotation speed and gear ratio are such that the slippage of the clutch is reduced to zero. During resynchronization, disk 1 must quickly obtain a new rotation speed, which leads to a significant inertial torque. Therefore, the output torque of the shaft 47 during the resynchronization process is controlled by slipping of the clutch, regardless of the displacement pressure in the cylinder 50, and can be reduced, as described above, to ensure the displacement and change of the speed of rotation of the disk 1 to the optimal value. Resynchronization must be quick to minimize clutch slippage in many applications. Therefore, the disk 1 is made as light as possible, and the shape is selected taking into account the minimization of inertia, in order to provide a resynchronization time of less than 0.1 seconds at full speed. For example, in a vehicle, disc 1 can be made of high quality bearing steel to be lightweight, and using a single disc can minimize inertia. It should also be noted that at low rotational speeds during the resynchronization process large axial displacements can occur between the disk 1 and the cones 7, which can be compensated by the above-described procedure for adjusting the bias pressure. However, at sufficiently high speeds and due to the physical characteristics, significant axial displacement does not occur.

In order to quickly displace the cylinder 50 during the resynchronization process, it is desirable to install control bypass valves 81 and 82 in order to minimize resistance to fluid flow. Valves 81 and 82 are connected, as shown in the drawing, and are installed as close as possible to the cylinder 50, allowing as much fluid flow as possible. Valve 81 is designed to allow fluid to pass through when the cylinder 50 extends quickly, and valve 82 is designed to allow fluid to pass through when the cylinder 50 is quickly retracted. Note that the pilot line of valve 81 is connected to line 73, and the pilot line of valve 82 is connected to line 72. Control valves are installed in series with each of the valves 81 and 82 to block back flow. Valves 81 and 82 may be arranged to proportionally open and allow fluid to pass through whenever the corresponding pilot pressure on them becomes negative. Thus, when the cylinder 50 is rapidly retracted during resynchronization upward, the flow resistance causes a negative pressure in line 72, so that the valve 82 is proportionally opened, allowing fluid to pass from the back to the front of the cylinder 50 with proportional resistance.

Similarly, when the cylinder 50 extends rapidly during the resynchronization process, the resistance of the fluid flow causes a negative pressure in line 73, so that the valve 81 is proportionally opened, passing fluid from the front to the rear of the cylinder 50.

In order to minimize slipping of the clutch during the transition from one range of rotation speeds to another, the speed of movement of the cylinder 50 during resynchronization should be as high as possible, and at the same time, the disk 1 cannot instantly move to a new synchronization point due to inertia. Therefore, the displacement rate of the cylinder 50 must be adjusted during the resynchronization process in order to coordinate the acceleration / deceleration time of the disk 1 at each and at all possible rotation speeds and all values of the load moment. Note that the speed of movement of the cylinder 50 in this situation depends on the following factors: (1) the prevailing reaction of the torque (load) between the shaft 2 and the sleeve 5; (2) flow resistance through valves 81 and 82; (3) the magnitude of the decrease in bias pressure due to the pilot pressure through line 107. Since the valves 81 and 82 are proportional, their degree of opening will be proportional to the combination of the load moment and the decrease in bias pressure. Thus, at a given load moment, the displacement rate of the cylinder 50 can be controlled by adjusting the degree of reduction of the displacement pressure. Note that the pressure in lines 106 and 108 and in lines 107 and 109 is ensured by the pressure in line 85, which is proportional to the load moment (however, it cannot be lower than the set minimum value). Therefore, the rate of change of the bias pressure during the transition from one speed range to another should be proportional to the load moment and, therefore, the speed of the disk 1 during resynchronization is also proportional to the load moment and the acceleration / deceleration time is always consistent. When operating at high speed, higher speed values require a proportional increase in the acceleration / deceleration time of the disk 1 and, therefore, a lower speed of the cylinder 50. Thus, the speed increase signal through line 91 counteracts the pressure in line 85 in the valve 103, so that the speed of the disk 1 decreases inversely with the increase in speed. Therefore, the resynchronization speed of the disk 1 during switching from one range to another decreases inversely with the rotation speed and the acceleration / deceleration time always turns out to be consistent for all permissible values of the rotation speed and load moment. Accordingly, the fastest resynchronization of the disk 1 occurs at the minimum rotation speed and maximum load torque, and the slowest resynchronization occurs at the maximum rotation speed and minimum load torque. Note that at maximum rotation speed and maximum load torque, the resynchronization time should be less than 0.1 seconds. It should also be noted with reference to FIG. 3 that, whenever the resynchronization process starts, the lever 6 immediately leaves contact with the engine 97, allowing the valve 63 to turn off. This starts the reset of the valve 103 and prevents the valve 87 from going beyond the boundaries. It is advisable to choose a reset time of the valve 103 corresponding to the resynchronization time of the disk 1 at each and all rotation speeds and all values of the load moment. Thus, by using the valve 103, the pressure in line 85 reduces the discharge time of the valve 103 with a proportional increase in the load moment. In addition, the speed signal in line 91 increases the reset time of the valve 103 with a proportional increase in the input rotation speed. Therefore, the reset of the valve 103 occurs exactly upon completion of the resynchronization of the disk 1 at each and all values of the input speeds and load moments. After the valve 103 is released, the pressure in the lines 105, 106 and 108 is relieved, so that the clutch engages fully (without slipping) and the displacement pressure of the cylinder 50 is restored to a value equivalent to the torque. With this range of speeds of rotation, the torque is controlled by the bias pressure applied to the cylinder 50. However, during the switch from one speed range to another and, in particular, during the resynchronization of the disk 1, which occurs relatively quickly, the torque is temporarily regulated by clutch slip.

When the acceleration of rotation of the shaft 47 continues in the second speed range, the disk 1 again shifts to the wide end of the cones 7, whereby switching to the third speed range occurs. In the third range, the clutch 41 remains engaged, the clutch 32 engages, the clutch 39 disengages, and the clutches 33 and 39 rotate freely. In this embodiment, as an example, planet gear sets 36 and 42 can provide a two to one (2: 1) gear ratio between the sun gear 35 and the ring gear 43, so that the ring gear 44 rotates in the forward direction at half speed in the third range of working speeds. The valves react, and the disk 1 again moves in the third speed range to the wide ends of the cones 7. When the disk 1 again reaches the wide ends of the cones 7, a switch to the fourth speed range occurs in the same way. In the fourth range, the clutches 32 and 39 are engaged and the clutches 33 and 41 are in free rotation, so that the ring gear 44 rotates in the forward direction at full speed in the fourth speed range. The full range of speeds for the vehicle is provided by these four ranges, however, other variants of ranges and gear ratios are considered. Thus, when the disk 1 again reaches the wide ends of the cones 7 in the fourth speed range, switching does not occur, since the maximum speed is reached.

Further, the device in a preferred embodiment operates as follows. At any speed and load moment on the shaft 47, when the actuator 88 returns to the neutral position, the signal on line 89 is turned off, so that the module 59 sets the signal on line 65 to zero, which sets the torque to zero, regardless of the speed signals in lines 91 and 93.

To provide negative torque and deceleration of the shaft 47, the actuator 88 moves past the neutral position in the deceleration direction, thereby piloting the pressure in the module 58 is switched from line 77 to line 76, and the valve 69 is turned on in the opposite direction. Thus, in order to provide a negative output torque, a bias pressure is applied to the cylinder 50 via line 72, so that the disk 1 moves proportionally in the direction of the narrow ends of the cones 7. As a result of the deceleration of the sun gear 49, negative torque is transmitted through the planet gears 45 to the sleeve 46 and onto the shaft 47. A slight movement of the actuator 88 to the deceleration position also leads to the direction of the signal through line 95 to module 94, which proportionally reduces the signal on line 93, to ory in turn causes the maximum deceleration of the motor. When the engine speed is slowed down, the signal 91 likewise decreases, however, the already reduced signal in line 93 causes the module 59 to transmit a torque increase signal to line 65. Thus, the engine speed can increase significantly with a significant negative moment applied to load, even if the module 94 provides a signal to reduce the speed of the engine. To prevent overspeeding of the engine, module 59 is calibrated so that the maximum engine speed signal on line 91 causes the module 59 to cancel the engine speed increase signal on line 65 and thus reduce the negative load moment. Further movement of the actuator 88 to the deceleration position proportionally increases the negative torque, but allows the engine shutter to be in the open position.

Note that the downshift operation of the clutch occurs when the lever 6 moves to the 6-D position in a similar manner, but in the opposite direction than when the upshift occurs and the lever 6 moves to the 6-U position. Thus, while the shaft 47 slows down the rotation, the speed ranges are switched downwards, in the reverse order to the above ranges up, until the lower range is reached again.

To provide reverse rotation of the output shaft 47, which is possible only in the lower speed range, the actuator or pedal 88 is moved to the deceleration position, thereby the valve 69 is switched in the opposite direction and the cylinder 50 applies a downward movement force to the disk 1. When the pedal 88 is further moved to the deceleration position, module 94 provides a signal for increasing the engine speed, so that the speed signal in line 91 causes the module 59 to output a signal proportional to the torque in line 65. Thus Thus, the disk 1 moves, bypassing the synchronization point in the direction of the narrow ends of the cones 7 to reverse the direction of rotation of the output shaft 47.

When moving forward at a sufficient speed, a downward shift with a positive torque can occur, as in the case of overload on the shaft 47. As with all downshifts, when the valve 103 is set to the downshift position, pressure is applied to line 108 instead of line 106. If positive torque valve 80 does not switch in the opposite direction, so line 108 runs in parallel with line 109. Thus, valve 64 does not immediately transmit a pulse and valve 87 does not immediately select a new clutch combination Eden. The pressure through line 105 leads to a decrease in the pressure of the clutch, while the pressure through line 109 to the pressure relief valve 60 leads to an increase in bias pressure on the cylinder 50. Thus, the active clutches slip at the prevailing load moment, allowing the disk 1 to move in the direction of the wide ends of the cones 7 with increasing force on the cylinder 50. The valve 103 is released, so that the pressure is removed from the line 109, leading to the issuance of a delayed pulse from the valve 64 to the valve 87, so that the next lower comb tion clutch is engaged precisely when disk 1 reaches a new synchronization point. The clutches are then restored to full (without slipping) engagement, and the cylinder 50 restores the pressure equivalent to the torque. Thus, switching down with a positive torque on the load is provided without changing the torque on the shaft 47.

Likewise, upward switching occurs with negative torque, except that the delayed engagement of the clutch ends in a different way. Note that in the case of negative torque, the valves 69 and 80 switch to the reverse position. In addition, since the valve 103 is set to the up switch position, the pressure is directed to the line 106 instead of the line 108, however, the valve 80 is switched in the opposite direction, so that the pressure is sent to the line 109. Thus, the delayed step pulse is transmitted from the valve 64 to the valve 87, so that a new combination of clutches engages, full engagement of the clutch is restored, and pressure equivalent to torsion moment is restored in cylinder 50. All this happens simultaneously with reaching disk 1 new synchronization position. Thus, switching up or switching down with positive or negative torque can be achieved without changing or losing the output torque at any permissible load moment.

Thus, from the above description of the operation of the preferred embodiment of the device, it follows that the present invention provides a mechanical power transmission device with stepless variation of the gear ratio and high energy intensity. The invention provides the achievement of the described objectives and advantages, while being effective in relation to the costs of production and use in various applications, including vehicle transmissions.

It should also be noted that the invention is not limited to the described particular embodiments, various changes and additions are possible within the scope of the invention. For example, a lubrication system may provide cooling and lubrication of bearings, gears, shafts, etc. In addition, various embodiments of the invention may include a different number of conical shafts, different rotation speeds and gear ratios, a different number of gears, different control methods, depending on the specific application.

Claims (27)

1. A device for transmitting power with a change in the gear ratio of the speeds of the output part relative to the drive input part, comprising input and output parts mounted in a housing rotatably, the input part being in drive connection with at least one driving traction part for transmitting torque moment to said at least one leading traction part, a disk part located in the joint with a rotatably mounted shaft coaxially and in the drive connection and with at least one driving traction part, this disk part is a driven at least one driving traction part for transmitting torque to a rotating shaft, the disk being selectively axially movable on a rotatably mounted shaft, at least partially in response to the torque applied to the output part to reposition the disk part relative to at least one driving traction part to selectively change the traction gear sheniya traction between the driving part and the drive, wherein the output element selectively engages a rotatably mounted shaft for transmitting torque thereby at a certain transmission ratio of speeds relative to the input part. 2. The device for power transmission according to claim 1, characterized in that the plurality of leading traction parts are located on the periphery of the disk part, each of the leading traction parts contains a conical rotor, the edge of which is located almost parallel and is in traction mesh with the disk part. 3. The device for transmitting power according to claim 1, characterized in that it further comprises a shear mechanism associated with the disk part for selectively moving the disk part to a predetermined axial position on a rotatably mounted shaft relative to at least one driving traction parts for changing the output torque transmitted to said rotatably mounted shaft. 4. The device for transmitting power according to claim 3, characterized in that the shear mechanism comprises a helical groove made in a rotatably mounted shaft. 5. The device for power transmission according to claim 1, characterized in that at least one of said driving traction part comprises a clamping mechanism designed to act on at least one leading traction part to selectively ensure effective traction contact of this leading traction part with disc part. 6. The device for transmitting power according to claim 1, characterized in that at least one leading traction part is a rotor containing a shaft, which at one end contains a gear mating and driven with the input part, and at the other end contains the output gear, and the output gear forms a drive connection with the crown gear of the first set of planetary gears, and the set of planetary gears also includes a sun gear and a set of gears installed in the carrier, while the set of planetary gears finds I selectively engaged with the output element to transmit torque. 7. The power transmission device according to claim 6, characterized in that the rotating shaft is a driven disk part and forms a drive connection with the output sun gear connected by the drive connection to the output part, and the set of planetary gears is in a selective drive connection with the sun gear so This provides a parallel differential drive system through which the disk part and at least one leading traction part transmit torque to the output part with a certain datochnym velocity ratio. 8. The power transmission device according to claim 1, characterized in that it further comprises a differential gear unit operable in connection with an input power source, wherein the differential gear unit comprises a ring gear selectively engaged to drive a set of output planetary gears designed to selectively drive the output shaft. 9. The device for transmitting power according to claim 8, characterized in that the ratio of the output torque transmitted to the output part from the rotating shaft or from the differential gear unit, which are both driven by at least one leading traction part, is consistent the ratio synchronization point based on the required output torque from the output part due to the axial movement of the disk part. 10. The device for transmitting power according to claim 8, characterized in that it further comprises a control system associated with a differential gear unit, the control system receives control signals relating at least to the input and output transmission speeds, while the control system interacts with a differential gear unit to provide the required output speed and torque, selected to match the output power of the gearbox with the input power of the power source. 11. The device for transmitting power according to claim 8, characterized in that it further comprises a control system associated with the differential gear unit and the disk part, the control system responding to control signals relative to the axial position of the disk part, while the control system interacts with the differential gear unit to provide the required selected gear ratios of the speeds between the disk part and the differential gear unit. 12. The device for transmitting power according to claim 1, characterized in that the control system responds, at least in part, to the control signals related to the axial position of the disk part. 13. A device for transmitting power, comprising input and output parts mounted in a housing rotatably, the input part being in drive connection with at least one driving traction part for transmitting torque to this at least one driving traction parts, a disk part mounted in conjunction with a rotating shaft coaxially and in a drive connection with the aforementioned at least one leading traction part, wherein the disk part is driven by at least one leading traction an assembly for transmitting torque to a rotating shaft, wherein the disk can selectively move axially on the rotating shaft to change the position of the disk part relative to at least one leading traction part to selectively change the traction gear ratio between the leading traction part and the disk, while the output part is installed with the possibility of selective engagement with a rotating shaft to thereby transmit torque at a predetermined gear from wearing speeds relative to the input part, and at least one leading traction part is a rotor having a mounted shaft mounted in the bearing mechanism, and the clamping mechanism is a piston unit mounted together with the bearing, the piston is mounted with the possibility of selective application of axial load to the rotor through the shaft. 14. A device for transmitting power, comprising input and output parts mounted in a housing rotatably, the input part being in drive connection with at least one driving traction part for transmitting torque to at least one driving traction part , a disk part mounted in articulation with a rotating shaft coaxially and in a drive connection with at least one leading traction part, the disk part being a driven at least one leading traction part for I transmitting torque to the rotating shaft, the disk being selectively axially movable on the rotating shaft to change the position of the disk part relative to at least one driving traction part to selectively change the traction gear ratio between the driving traction part and the disk, the output part selectively engages with a rotating shaft to thereby transmit torque at a predetermined gear ratio of speeds relative to the input part, and dis the detail contains a sleeve mounted on a rotating shaft, the sleeve is torsionally connected to the rotating shaft by a screw groove made in the rotating shaft. 15. A continuously variable transmission comprising a disk part mounted in conjunction with the rotating shaft coaxially, the disk part being in drive connection with a plurality of driving traction parts, the plurality of driving traction parts including a leading surface parallel to the axis of the disk part, the disk part being mounted selective axial movement on a rotating shaft to change the position of the disk part relative to the set of leading traction parts for selective and To change the traction gear ratio between a plurality of driving traction parts and a disk part, a plurality of driving traction parts are set so as to be driven by an input power source from one end thereof, the second end being in a drive connection with a gear system comprising a ring gear in the drive connected to a set of output planetary gears, while the rotating shaft is mounted to drive the output sun gear associated with the set of planetary gears ter, while the sun gear is also in the drive connection with a set of planetary gears to provide a parallel differential drive through which the disk part and a plurality of leading traction parts transmit the output torque to the output part connected in the drive connection to the planet gears. 16. The continuously variable transmission according to claim 15, characterized in that it further comprises a shear mechanism for controlling the axial position of the disk part relative to the plurality of driving traction parts, whereby the gear ratio between the disk part and the driving traction parts is changed in order to control the output speed and torque output shaft. 17. The continuously variable transmission according to claim 16, characterized in that the shear mechanism comprises a helical groove made in a rotatably mounted shaft. 18. The continuously variable transmission according to claim 15, wherein the gear system is a selectable multi-band gear system. 19. The continuously variable transmission according to claim 15, characterized in that it further comprises a control system associated with the gear system and the disk part, while the control system responds to control signals, which at least partially relate to the axial position of the disk part, the control system interacts with the gear system to provide the required selected gear ratios of the speeds between the disk part and the gear system. 20. A continuously variable transmission comprising an input power source coupled in a drive connection to at least one driving traction part, this at least one driving traction part has a leading surface, a disk part driven by at least one leading traction part, and this disk part is torsion connected with the coaxial shaft and mounted with the possibility of selective movement relative to the leading surface to change the gear ratio of the torque transmitted to it from at least e, of one driving traction part, a differential gear unit coupled to an input power source, wherein the differential gear unit comprises a ring gear selectively engaged to drive a set of output planet gears intended to selectively drive the output sun gear, shaft, the coaxial shaft is designed to transmit torque to the output sun gear with a predetermined gear ratio together with the differential gear unit, and the position of the aforementioned disk parts relative to the leading surface is selectively adjusted to change the gear ratio of torque and speed transmitted to the sun gear by said coaxial shaft to the differential gear unit. 21. The continuously variable transmission according to claim 20, characterized in that it further comprises a shear mechanism for controlling the axial position of the disk part relative to the leading surface. 22. The continuously variable transmission according to claim 21, characterized in that it further comprises a control system associated with the shear mechanism and a differential gear unit, a control system controls the shear mechanism for selectively changing the position of the disk part, this control system also receives control signals relative to at least the input and output transmission rates, the control system controls the matching of the transmission output power with the input power source. 23. The continuously variable transmission according to item 21, wherein the shear mechanism comprises a helical groove made in a coaxial shaft. 24. The continuously variable transmission according to claim 20, characterized in that the position of the disk part relative to the driving surface is changed in response to the output torque to selectively maintain a synchronized relationship between speed and torque transmitted to the sun gear by the coaxial shaft relative to the differential gear unit. 25. The continuously variable transmission according to claim 20, characterized in that the differential gear block comprises a gear system comprising at least two sets of planetary gears providing a predetermined gear ratio to provide the desired speed range at the output. 26. The continuously variable transmission according to claim 20, characterized in that at least one leading traction part comprises a clamp mechanism designed to act on at least one leading traction part to selectively clamp this at least one leading traction part in order to ensure effective traction contact with the disk part. 27. The continuously variable transmission according to claim 20, characterized in that it further comprises a control system associated with the differential gear unit and the disk part, the control system responding to control signals relative to the axial position of the disk part, while the control system interacts with the differential gear unit to provide the required selected gear ratios of the speeds between the disk part and the differential gear block.

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